Torque biased friction hinge for a tensioner

ABSTRACT

A pivot joint for a single blade spring tensioner or double blade spring tensioner generates a biased friction loss at a pivot joint. The friction torque generated in one direction of joint rotation is greater than the friction torque generated in the opposite direction of rotation. In one embodiment, the direction of torque bias at the joints is used to make the damping of the blade spring greater in compression than in extension. High damping in compression and low damping in extension is a desirable characteristic for a tensioner in most applications. The device may be designed to act as a one-way clutch or as a two-way clutch. The device may also be designed to apply large friction torques only when the joint undergoes large amplitude motions (loss of control).

REFERENCE TO RELATED APPLICATIONS

This application claims one or more inventions which were disclosed inProvisional Application No. 60/765,777, filed Feb. 7, 2006, entitled“TORQUE BIASED FRICTION HINGE FOR A TENSIONER”, Provisional ApplicationNo. 60/822,520, filed Aug. 16, 2006, entitled “SELF-ENERGIZING BRAKE FORA TENSIONER”, and Provisional Application No. 60/863,815, filed Nov. 1,2006, entitled “BLADE TENSIONER WITH OPPOSING SPANS”. The benefit under35 USC §119(e) of the United States provisional applications is herebyclaimed, and the aforementioned applications are hereby incorporatedherein by reference.

This is a continuation-in-part patent application of copending PCTapplication number PCT/US2007/060941, filed Jan. 24, 2007, entitled“BLADE TENSIONER WITH OPPOSING SPANS”. The aforementioned application ishereby incorporated herein by reference.

This is a continuation-in-part patent application of copending PCTapplication number PCT/US2007/060945, filed Jan. 24, 2007, entitled“SELF-ENERGIZING BRAKE FOR A TENSIONER”. The aforementioned applicationis hereby incorporated herein by reference.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The invention pertains to the field of tensioners. More particularly,the invention pertains to a torque biased friction hinge for atensioner.

2. Description of Related Art

A blade-type tensioner imparts tension on a chain. A blade-typetensioner generally includes a plastic blade shoe with an arcuatelycurved chain sliding face, one or more leaf spring-shaped blade springsopposite the chain sliding face, and a metal base that swingablysupports a proximal end portion of the blade shoe and slidably supportsa distal end portion of the blade shoe.

During operation, a chain slides and travels along the chain slidingface of the blade shoe. A resilient force due to the elastic deformationof the blade shoe and the blade spring is applied to the chain throughthe blade shoe to maintain proper chain tension.

SUMMARY OF THE INVENTION

A pivot joint for a single blade spring tensioner or double blade springtensioner generates a biased friction loss at a pivot joint. Thefriction torque generated in one direction of joint rotation is greaterthan the friction torque generated in the opposite direction ofrotation. In one embodiment, the direction of torque bias at the jointsis used to make the damping of the blade spring greater in compressionthan in extension. High damping in compression and low damping inextension is a desirable characteristic for a tensioner in mostapplications. The device may be designed to act as a one-way clutch oras a two-way clutch. The device may also be designed to apply largefriction torques only when the joint undergoes large amplitude motions(loss of control).

A pivot joint of the present invention includes a pressure plate havinga pressure plate surface, a clutch plate, a pivot arm, a pivot pin, andfirst and second springs. A clutch plate hole extends through the clutchplate. A clutch plate surface is in contact with the pressure platesurface. The clutch plate has at least one inclined clutch contact areaopposite the clutch plate surface. The pivot aim has a pivot hole andincludes at least one inclined arm contact area in contact with andcomplementary in shape to the inclined clutch contact area. The pivotpin includes a pivot pin head and a pivot pin shaft extending throughthe pivot hole and the clutch plate hole and into the pressure plate.

The first spring is mounted on the pivot pin between the pivot pin headand the pivot arm to urge the pivot arm away from the head of the pivotpin and to provide a first spring preload on the pivot arm and theclutch plate. The spring also provides an end gap between the pivot pinhead and the pivot arm to be taken up upon tensioning by an increase ina stack height of the clutch plate and the pivot arm around the pivotpin. When the end gap has not been taken up and the pivot arm rotates ina compression direction, the inclined arm contact area rotates aroundthe pivot pin relative to the inclined clutch contact area. Thus thedistance between the inclined arm contact area and the pressure plateand the stack height around the pivot pin are increased.

In a first embodiment, a second spring is mounted on the pivot pin. Thesecond spring acts on a surface of the clutch plate to bias the clutchplate toward the pressure plate for providing a second spring preload onthe clutch plate. Preferably, a coefficient of friction between theclutch plate and the pressure plate, a coefficient of friction betweenthe inclined clutch contact area and the inclined arm contact area, anangle of inclination of the inclined arm contact area with respect to aplane parallel to the clutch plate surface, the first spring preload,the second spring preload, an active radius of the inclined arm contactarea, and an active radius of the clutch plate are chosen such that inpivot arm compression a torque to slip the inclined plane contact areabecomes greater than a torque to slip the clutch plate above a criticaltorque that occurs at a pivot arm displacement greater than adisplacement required to take up the end gap. Preferably, a coefficientof friction between the clutch plate and the pressure plate, acoefficient of friction between the inclined clutch contact area and theinclined arm contact area, an angle of inclination of the inclined armcontact area with respect to a plane parallel to the clutch platesurface, the first spring preload, the second spring preload, an activeradius of the inclined arm contact area, and an active radius of theclutch plate are chosen such that in pivot arm extension a torque toslip the inclined plane contact area is less than a torque to slip theclutch plate. In one embodiment, a frictional material is located on theclutch plate surface or the pressure plate surface. In anotherembodiment, the pressure plate surface is formed to change the torquerequired for rotation of the clutch plate with respect to the pressureplate.

The first spring is preferably a Belleville spring. The inclined clutchcontact area is preferably centered about the pivot hole and has a sloperunning tangential to a constant radius from a center point of the pivothole. In one embodiment, the pressure plate is mounted to a stationarysurface to prevent rotation of the pressure plate relative to thesurface.

In one embodiment, the second spring acts on a surface of the pivot armto bias the clutch plate toward the pressure plate. In anotherembodiment, the second spring acts on a surface of the pivot pin to biasthe clutch plate toward the pressure plate.

Preferably, a coefficient of friction between the clutch plate and thepressure plate (μ_(C)), a coefficient of friction between the inclinedclutch contact area and the inclined arm contact area (μ_(R)), an angleof inclination of the inclined arm contact area with respect to a planeparallel to the clutch plate surface (θ), the first spring preload(F_(S1)), the second spring preload (F_(S2)), an active radius of theinclined arm contact area (R_(R)), and an active radius of the clutchplate (R_(C)) are chosen such that:

$\frac{F_{S\; 2}}{F_{S\; 1}} > {\frac{{R_{R}\left( {{\sin\;\theta} + {\mu_{R}\cos\;\theta}} \right)} - {R_{C}{\mu_{C}\left( {{\cos\;\theta} - {\mu_{R}\sin\;\theta}} \right)}}}{R_{C}{\mu_{C}\left( {{\cos\;\theta} - {\mu_{R}\sin\;\theta}} \right)}}.}$

A tensioner of the present invention is also disclosed. In oneembodiment, at least one blade spring is mounted in the tensioner arm,and the distal end portion of the tensioner arm is slidingly received ona sliding surface.

In another embodiment, the distal end portion of the tensioner arm ispivotally attached to a second pivot pin. The tensioner further includesat least one blade spring mounted in the tensioner arm. A secondtensioner arm has a proximal end portion pivotally attached to the pivotpin and a distal end portion pivotally attached to the second pivot pin.A wear surface on the second tensioner arm is slidingly received on asliding surface. The second tensioner arm may serve as the pressureplate for the joint.

In yet another embodiment of the present invention, the pivot jointincludes a pressure plate, a clutch plate, a pivot arm, a pivot pin, anda preload spring. The pressure plate has a pressure plate surface. Theclutch plate has a clutch plate hole extending through the clutch plateand a clutch plate surface in contact with the pressure plate surface.The clutch plate also has at least one inclined clutch contact areaopposite the clutch plate surface. The pivot arm has a pivot hole andincludes at least one inclined arm contact area facing and complementaryin shape to the inclined clutch contact area. The pivot pin includes apivot pin head and a pivot pin shaft extending through the pivot holeand the clutch plate hole and into the pressure plate. The preloadspring is mounted on the pivot pin between the pivot arm and the clutchplate. The preload spring urges the pivot arm away from the clutch plateand provides a spring preload on the clutch plate. The preload springalso provides a preload gap between the inclined clutch contact area andthe inclined arm contact area to be taken up by an increase in a stackheight of the clutch plate and the pivot arm around the pivot pin.

In another embodiment of the present invention, the tensioner includes apressure plate, a clutch plate, a tensioner arm, a pivot pin, and apreload spring. The pressure plate has a pressure plate surface. Theclutch plate has a clutch plate hole extending through the clutch plateand a clutch plate surface in contact with the pressure plate surface.The clutch plate also has at least one inclined clutch contact areaopposite the clutch plate surface. The tensioner arm for tensioning achain or belt has a pivot hole at a proximal end portion. The tensionerarm includes at least one inclined arm contact area facing andcomplementary in shape to the inclined clutch contact area. The pivotpin includes a pivot pin head and a pivot pin shaft extending throughthe pivot hole and the clutch plate hole and into the pressure plate.The preload spring is mounted on the pivot pin between the pivot arm andthe clutch plate and urges the pivot arm away from the clutch plate. Thepreload spring provides a spring preload on the clutch plate. Thepreload spring also provides a preload gap between the inclined clutchcontact area and the inclined arm contact area to be taken up by anincrease in a stack height of the clutch plate and the pivot arm aroundthe pivot pin.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a single blade spring tensioner in an embodiment of thepresent invention.

FIG. 2A shows a partial cross-sectional view along lines 2A-2A of FIG.1.

FIG. 2B shows the tensioner of FIG. 2A in a high frictional lossesstate.

FIG. 3 shows a double blade spring tensioner in an embodiment of thepresent invention.

FIG. 4 shows a partial cross-sectional view along lines 4-4 of FIG. 3.

FIG. 5A shows two inclined plane contact areas in one embodiment of thepresent invention.

FIG. 5B shows a side view of FIG. 5A.

FIG. 6 shows three inclined plane contact areas in an alternativeembodiment of the present invention.

FIG. 7 shows four inclined plane contact areas in an alternativeembodiment of the present invention.

FIG. 8 shows a tensioner in an embodiment of the present invention.

FIG. 9 shows a partial cross-sectional view along lines 9-9 of FIG. 8.

FIG. 10 shows a flowchart of a method of the present invention, whichcreates low frictional losses.

FIG. 11 shows a flowchart of a method of the present invention, whichcreates high frictional losses.

FIG. 12 shows a free body diagram for compression of a friction hinge ofthe present invention.

FIG. 13 shows a free body diagram for extension of a friction hinge ofthe present invention.

FIG. 14 shows a torque-displacement curve for a friction hinge of thepresent invention.

FIG. 15 shows a single-spring two-way clutch in an embodiment of thepresent invention.

FIG. 16 shows a one-way clutch tensioner in an embodiment of the presentinvention.

FIG. 17 shows a partial cross-sectional view along lines 17-17 of FIG.16.

DETAILED DESCRIPTION OF THE INVENTION

In a first set of embodiments, a friction hinge of the present inventionacts as a two-way clutch. The friction hinge includes a first springbiasing the pivot arm and the clutch toward the pressure plate and asecond spring biasing only the clutch toward the pressure plate. For thedouble spring device, the clutch may slip for an infinite displacementin the high torque direction (compression) at the critical torque value.The double spring device may also slip for infinite displacements in thelow torque direction (extension), because the back faces of the rampscollide and allow the clutch to be driven.

In a second set of embodiments, a friction hinge of the presentinvention acts as a one-way clutch. The friction hinge includes a singlespring biasing the pivot arm and the clutch toward the pressure plate.The single spring device has a limited possible displacement in the hightorque direction (compression), because the clutch never slips in thatdirection. The single spring device may slip for infinite displacementsin the low torque direction (extension), because the back faces of theramps collide and allow the clutch to be driven. The rotation in thehigh torque direction required for the single spring device to bind thepin and generate enough torque to prevent rotation serves as thebacklash for the one way clutch device.

Tensioner compression is defined herein as a rotation of the tensionerarm in a direction that reduces chain tension, whereas tensionerextension is defined herein as a rotation of the tensioner arm in adirection that increases chain tension.

In this application, a single blade spring tensioner refers to atensioner having a single flexible element used to tension a chain drivewith one end fixed to ground by a pivot and the other end sliding on aramp that is fixed to ground. The flexible element preferably has apolymer wear surface and one or more metallic blade springs, althoughother materials could alternatively be used.

A double blade spring tensioner refers to a tensioner having twoopposing flexible elements joined together at each end by pin joints.One of the pin joints allows both elements to rotate relative to ground.The other pin joint is not constrained to ground. In a double bladespring tensioner, one of the wear surfaces contacts the chain, and theother wear surface contacts a ramp that is fixed to ground. The flexibleelements in either tensioner include a wear surface connected to one ormore blade springs. In a preferred embodiment, the wear surfaces aremade of a polymer and the blade springs are metallic, however, othermaterials known in the art could alternatively be used. In anunconstrained state, each flexible element approximates a constantradius.

In this application, the term “ground” represents a stationary referenceframe for the chain drive. Common examples of ground include, but arenot limited to, an engine block, a mounting bracket, a transmissioncase, and a machine frame or case.

Although a friction hinge of the present invention is described for usein a blade tensioner, the present invention is applicable to a pivot forany tensioning arm for tensioning a chain or a belt.

A pivot joint for a single blade spring tensioner or double blade springtensioner generates a biased friction loss at a pivot joint. Thefriction torque generated in one direction of joint rotation is greaterthan the friction torque generated in the opposite direction ofrotation. In one embodiment, the direction of torque bias at the jointsis used to make the damping of the blade spring greater in compressionthan in extension. High damping in compression and low damping inextension is a desirable characteristic for a tensioner in mostapplications. The device may also be designed to apply large frictionforces only when the joint undergoes large amplitude motions (loss ofcontrol).

The friction joint of the present invention may be used at the pivot endof a single blade spring tensioner or at either pivot on a double bladespring tensioner. In one embodiment, a mating set of inclined planes arearranged on two circular contact areas, however, any number of inclinedplanes may be used in the present invention. More planes may be used todecrease wear and decrease contact pressures on the hinge. The inclinedplanes are preferably made of a polymer by injection molding or from apowdered metal (PM). Thus, there is little, if any, cost difference formanufacture with different numbers of inclined planes, but a specificnumber of inclined planes may be preferable for a specific application.The inclined plane contact areas are centered about a pivot hole in ablade spring flexible element, and the inclined planes in the contactarea are arranged with the slope of each plane running tangential to aconstant radius from the center point of the pivot hole.

In one embodiment, the inclined plane contact area is preferably apolymer material that is formed during the molding process for the bladetensioner polymer wear surface. In an alternative embodiment, theinclined plane contact area is a metallic material fixed to thetensioner flexible element, and the flexible element is preferablymolded directly around the metallic material, which is preferably apowdered metal part. The powdered metal part preferably has spurs aroundwhich the flexible element is molded to support torque transfer betweenthe inclined plane contact area and the flexible element.

The second inclined plane contact area is preferably identical andcomplementary in shape to the first contact area, but the second contactarea is preferably on a small cylindrical part that is referred to asthe clutch plate herein. The clutch plate is preferably made of apolymer material or a metallic material. In one embodiment, the clutchplate is made from powdered metal (PM) to eliminate the need to performcomplicated machining to produce the inclined planes on a metallic part.The clutch plate has a contact area on the side of the cylinder oppositethe inclined plane contact area, and this contact area is referred to asthe clutch contact area (or clutch plate contact area) herein. Theclutch contact area optionally has friction material bonded to thesurface. Friction materials include, but are not limited to, paper-basedwet clutch materials and brass. The clutch contact area is alsooptionally textured to alter the frictional properties, as is commonindustry practice with clutches. For example, the contact areapreferably has cross-hatched cuts or grooves in the contact areasurfaces of the lubricated parts.

The clutch plate mates with a pressure plate that is preferably eithermounted to ground or to another blade spring flexible element (in ajoint on a double blade tensioner). The joint is held together with arigid metallic pin that is long enough to provide a small end clearancein the joint when the pair of inclined planes is fully compressed. Thejoint end play is preferably sized so that a small relative rotation ofthe pair of inclined planes takes up the end gap and forces the stack ofcomponents to contact the end stops on the pin and directly load the pinin tension. A small relative rotation is consistent with a small aimdeflection of one to two millimeters or less, but is dependent on theapplication.

In the two-spring embodiments, a first preload spring and a secondpreload spring are preferably placed in the column of components tomaintain a nominal preload on the inclined planes and the clutch plate,when the column does not bind directly against the joint pin. The firstpreload spring urges the tensioner arm and the clutch plate toward thepressure plate. The first preload spring is preferably a Bellevillespring. A coil spring or any other device that would place a nominalpreload on the frictional plate may alternatively be used. A torsionspring may also be used. A washer is also optionally included with thespring. The second preload spring, preferably a coil spring, urges onlythe clutch plate toward the pressure plate.

The two preload springs place an initial friction force on the clutch,which is higher than the initial friction force on the inclined planes.The two preloads, in combination with the coefficient of frictionbetween the clutch plate and the pressure plate, the coefficient offriction between the mating inclined plane surfaces, the angle of theinclined planes, and the active radii of the clutch and the inclinedplanes are preferably selected to meet the following conditions duringtensioner compression. At low friction torques, the inclined planes slipbefore the clutch. At a critical torque, the friction torque on theclutch equals the friction torque on the inclined planes. Above thecritical torque, the clutch plate slips. This clutch slipping featureprevents part failure at high compression forces and is used to adjustthe coulomb damping of the tensioner.

When the joint rotates in the direction where low frictional losses aredesired, the following events preferably occur to generate rotation inthe joint with relatively low friction. The joint turns in the directionwhere the relative motion of the pair of inclined planes reduces thetotal column height. The flat back sides to the inclined planes collide,causing the flexible element to drive the clutch plate in rotation aboutthe pin. The inclined planes are fully collapsed so that the only normalforce between the clutch plate and the pressure plate is the nominalpreload generated by the Belleville spring and the second spring.Friction exists in the contact between the clutch plate and the pressureplate, but the torque generated by the friction is relatively low due tothe low normal force. The pitch of the inclined planes and the frictionproperties of the clutch and inclined planes are preferably designed tominimize the tendency of the inclined planes to stick when the directionof rotation changes from high desired friction to low desired friction.

When the joint rotates in the direction where high frictional losses aredesired, the following events preferably occur to generate rotation inthe joint with relatively high friction. The joint turns in thedirection where the relative motion of the pair of inclined planesincreases the total column height. The friction joint is preferablydesigned such that the clutch slips when a desirable or predeterminedcritical compression force on the hinge is exceeded. When thecompression force on the hinge is less than the critical value, theinclined planes rotate and spread apart rather than the clutch platerotating relative to the pressure plate without spreading the pair ofinclined planes. After a small rotation, the inclined planes increasethe total column height until all joint end play is removed and thecolumn binds against the ends of the joint pin, loading the pin directlyin tension. Further rotation causes the inclined planes and the clutchto act as a self-energizing clutch where great normal forces aregenerated, placing the joint pin in tension and the part stack column incompression. The high normal forces caused by the self-energizing clutchplace a relatively high frictional torque on the joint.

The slope of the inclined planes and the size of the end gap arepreferably selected so that the clutch does not self-energize for smalljoint rotations when the blade spring device is functioning properly andproviding adequate tensioning force to control the drive. Underconditions where the blade spring can not control the chain drive, theblade spring deflections are large and the joint rotational motion alsoincreases in amplitude. The clutch self-energizes during bladecompression as a result of the large rotational motion at the joint.Properly sizing the end gap and inclined plane characteristics allowsthe high friction forces to be generated only during running conditionswhere the additional friction force is actually needed to control thedrive. In other words, small deflections are damped completely by theflexible tensioner aim alone without damping by the clutch, and theclutch only activates at large deflections. Creating high frictionforces only when needed reduces wear on all parts over the lifetime ofthe device.

In a two-spring friction hinge of the present invention, the followingconditions preferably exist. During compression prior to binding thepin, the required torque to slip the inclined planes is less than therequired torque to slip the clutch. For some torque range after the pinbinds, the required torque to slip the inclined planes continues to beless than the required torque to slip the clutch. At some point afterthe pin binds, the required torque for the inclined planes to slipbecomes greater than the required torque to slip the clutch. Thus, attorque loads below a critical value, the clutch holds and the inclinedplanes slide past each other. At the point of critical torque load, theclutch of the friction hinge begins to slip to relieve the torque loadand prevent part failure.

The angle of the inclined planes, the coefficients of friction for theinclined planes and the clutch, and the active radii of the inclinedplanes and the clutch are preferably selected to meet these conditions.

FIGS. 1, 2A, and 2B show a single blade tensioner (1) with a pivot joint(2) of the present invention. The single blade tensioner (1) includes anarcuately curved tensioner arm (50) with a distal end (51), a pivotingproximal end (3), and a chain sliding face (52). The proximal end (3) ofthe arm (50) is pivotally mounted to the ground (13) (i.e., the enginehousing or a bracket) by a pin (14). The distal end (51) slides on aramp (19) that is fixed to ground. Although the ramp (19) is shown as awedge in the figures, other shapes that allow the tensioner to slide onthe ramp are also encompassed by the present invention. The arm (50)contains a blade spring (53) for supporting the chain sliding face (52).Although only one blade spring (53) is shown in the figures, multipleblade springs (53), which are preferably metallic, are alternativelyincluded.

The friction joint (2) is used at the pivot end (3) of the single bladespring tensioner. FIGS. 1, 2A, and 2B show a mating set of inclinedplanes (4A), (4B), (5A), (5B) arranged on two semi-circular contactareas, although any number of inclined planes may be used with thesingle blade spring tensioner (1). FIGS. 5A and 5B show a top view and aside view of the inclined plane contact areas (6), (7) for a clutchplate (9) with two inclined planes (4A), (4B).

Both sets of inclined plane contact areas are centered about a pivothole (8). One set extends from a blade spring flexible element (18),while the second set extends from the clutch plate (9). The slope of theinclined planes (4A), (4B), (5A), (5B) runs tangential to a constantradius from the center point of the pivot hole (8). The dashed area (88)in FIG. 2A is preferably made of a continuous polymer or a metallicinsert molded into the polymer portion of the blade spring. The clutchplate (9) has a clutch contact area (10) on the side of the cylinderopposite the inclined plane contact areas. The clutch contact area (10)optionally has friction material (11) bonded to its surface.

The clutch plate (9) mates with a pressure plate (12). In thisembodiment, the pressure plate (12) is mounted to ground (13). The jointis held together with a rigid metallic pin (14) that is long enough toprovide a small end clearance in the joint when the pair of inclinedplanes (4), (5) is fully compressed. The joint end play is preferablysized so that a small relative rotation of the pair of inclined planes(4), (5) takes up the end gap (15) and forces the stack of components tocontact the end stops on the pin and directly load the pin in tension. ABelleville spring (16) is preferably included to maintain a nominalpreload on the inclined planes (4), (5) and the clutch plate (9) whenthe column does not bind directly against the joint pin (14). A washer(17) is also optionally included. A coil spring (20) is preferablyincluded to maintain a nominal preload on the clutch plate (9), which isgreater than the preload on the inclined planes (4), (5). Thecoefficient of friction between the clutch plate (9) and the pressureplate (12) and the coefficient of friction between the mating inclinedplane surfaces (6), (7) together with the angle of inclination of theplanes, the spring preloads, and the active radii of the clutch and theinclined planes help to determine the performance of the friction hinge.

The single blade tensioner (1) is shown in a low frictional losses statein FIG. 2A and a high frictional losses state in FIG. 2B. In FIG. 2A,the end gap (15) is large and the inclined planes (4A), (4B), (5A), (5B)are aligned such that there is contact between complementary verticalsurfaces at the high ends of the inclined planes. In FIG. 2B, the stackheight has increased, and the end gap (15) has been taken up as rotationof the flexible element (18) inclined planes (5A), (5B) has caused thecomplementary inclined plane contact areas to come into frictionalcontact, forcing the flexible element (18) farther away from thepressure plate (12) flattening the Belleville spring (16), and bindingthe pivot pin (14). FIG. 2A and FIG. 2B are shown from the sameperspective of the flexible element (18), making it appear that theclutch plate (9) has rotated. It is important to note, however that itis actually the tensioner arm (50), and hence the flexible element (18),that rotates and the clutch plate (9) and pressure plate (12) remainstationary in going from the state of FIG. 2A to the state of FIG. 2B,unless there is slippage between the clutch plate (9) and the pressureplate (12), in which case the clutch plate (9) rotates as a result ofthe slippage.

The first preload spring (16) and the second preload spring (20) place anominal preload on the clutch plate (9), whereas only the first preloadspring (16) places a nominal preload on the inclined planes (4), (5).The preload on the clutch plate (9) is therefore greater than thepreload the inclined planes (4), (5). In the embodiment of FIG. 2A andFIG. 2B, the first preload spring (16) is shown as a Belleville spring.As shown in the embodiment of FIG. 4, a coil spring or other spring (46)may be used instead of the Belleville spring. In the embodiment of FIG.2A and FIG. 2B, the second preload spring (20) acts on the clutch plate(9) and the tensioner arm (18) to urge the clutch plate (9) away fromthe tensioner arm (18). As shown in the embodiment of FIG. 4, the secondpreload spring (49) may alternatively urge the clutch plate (39) awayfrom the pin (44) by contacting a surface of the pin.

The end gap (15) is typically only present in one direction, except ifthe rotation is small enough, and then the end gap is present in bothdirections. Consequently, the only normal forces applied to the clutchare from the Belleville spring (16) and the coil spring (20). Tensionerextension causes the inclined planes (4), (5) to spread, until the endgap (15) is taken up and contact occurs between the tensioner body andthe pin (14). This creates high normal forces and high friction torquesfrom the joint (2).

FIGS. 3 and 4 show a double blade tensioner (30) with the pivot joint(32) of the present invention. A tensioner arm (50) contains a bladespring (53) for supporting the chain sliding face (52). Although onlyone blade spring (53) is shown in each tensioner arm in the figures,multiple blade springs (53), which are preferably metallic, arealternatively included.

Both ends (33A), (33B) of the tensioner (30) are joined together by pinjoints. The distal pin joint (36) may be a prior art joint or a frictionjoint of the present invention. One of the wear surfaces is a chainsliding face (52) that contacts a chain, while the other wear surface(61) contacts a ramp (37) that is fixed to ground. Although the ramp(37) is shown as a wedge in the figures, other shapes that allow thetensioner to slide on the ramp are also encompassed by the presentinvention. The wear surface (61) slides along the ramp (37).

The friction joint (32) may be at either pivot end (33A), (33B) of thetensioner (30) but is shown at the right pivot end (33A) in the figures.A mating set of inclined planes (34A), (34B), (35A), (35B) is arrangedon two semi-circular contact areas, however, any number of inclinedplanes may be used with the double blade spring tensioner (30). Bothsets of inclined plane contact areas are centered about a pivot hole(38). One set extends from a blade spring flexible element (48A).Although the second set preferably extends from the clutch plate (39) asshown in FIG. 4, the second set may extend directly from the secondblade spring flexible element (48B) within the spirit of the presentinvention. The slope of the inclined planes (34A), (34B), (35A), (35B)runs tangential to a constant radius from the center point of the pivothole (38). The clutch contact area (40) optionally has friction material(41) bonded to the surface.

The clutch plate (39) mates with a pressure plate (42) that ispreferably mounted to the second blade spring flexible element (48B).The joint is held together with a pin (44), which is preferably a rigidmetallic pin that is long enough to provide a small end clearance in thejoint when the pair of inclined planes (34), (35) is fully compressed.The joint end play is preferably sized so that a small relative rotationof the pair of inclined planes takes up the end gap (45) and forces thestack of components to contact the end stops on the pin and directlyload the pin in tension. A first coil spring (46) maintains a nominalpreload on the inclined planes (34), (35) and the clutch plate (39) whenthe column does not bind directly against the joint pin (44). In analternative embodiment, a Belleville spring is used, similar to FIG. 2A.A washer (47) is also optionally included. A second spring (49) ispreferably included to maintain a nominal preload on the clutch plate(39), which is greater than the preload on the inclined planes (34),(35). In the embodiment of FIG. 4, the second spring (49) contacts theclutch plate (39) on one end and an expanded part of the pin (44) shaftat the other end. The coefficient of friction between the clutch plate(39) and the pressure plate (42) and the coefficient of friction betweenthe mating inclined plane surfaces, together with the angle ofinclination of the planes and the active radii of the clutch and theinclined planes help to determine the performance of the friction hinge.

In the double-blade embodiment, the flexible element (48B) is preferablyused as the pressure plate (42). The pressure plate section of theflexible element (48B) is preferably made of a polymer or a metallicsection molded into the flexible element (48B). The friction material(41) is preferably applied to the pressure plate (42) or the clutchplate (39).

It is not necessary that the right end (33A) of a double blade tensionerof the present invention be grounded. In a free end pivot embodiment ofa double blade tensioner, a pin retainer (31) located at the far end ofthe pin (44) is not grounded. In an embodiment with the pivot fixed toground, the retainer (31) represents ground. Additional frictionmaterial may optionally be placed between flexible element (48B) andground.

Although two inclined planes are shown in the figures discussed thusfar, any number of inclined planes may be used. For example, FIG. 6shows three inclined plane contact areas (70), (71), (72). Each of thecontact areas goes from a low incline (L) to a high incline (H). FIG. 7shows four inclined plane contact areas (80), (81), (82), (83). Each ofthe contact areas goes from a low incline (L) to a high incline (H). Asdiscussed above, more planes may be used to decrease wear and decreasecontact pressures on the hinge. A specific number of inclined planes maybe preferable for a specific application. Although a single pair ofinclined planes may be used, two or more are preferable so as not to puta moment on the inclined planes and create friction.

A torque-biased friction hinge of the present invention may be used withany type of tensioning arm. For example, FIGS. 8 and 9 show a pivotjoint (90) for a tensioner arm (91). The tensioner aim (91) is biasedtoward the tensioned device (92) by a tensioner arm biasing device (93).An alternate to the previously-described pivot pin is shown in thisembodiment. A dowel pin (94A) serves as the pin shaft and a snap ring(94B) placed around a narrowed portion of the dowel pin (94A) serves asthe pin head for this pivot pin. The dowel pin (94A) is preferablygrounded in a pressure plate (95). Stacked on the pin (94A) above thepressure plate (95) are a clutch plate (96), the tensioner arm (91), awasher (97), a spring (98) to bias the tensioner arm (91) toward theclutch plate (96) and provide an end gap (102), and the snap ring (94B).A dowel pin and a snap ring may be used interchangeably with a pin withan integral head in any of the embodiments described herein. Thetensioner arm (91) and the clutch plate (96) have complementary inclinedplanes (99A), (99B), (99C), (99D) forming complementary contactingsurfaces on the sides facing each other. Although only two sets ofinclined planes are shown in FIG. 9, any number of inclined planes maybe used in the present invention. A second spring (100) is preferablyincluded to maintain an additional nominal preload on the clutch plate(96).

FIGS. 10 and 11 show flowcharts of methods of the present invention.When the joint rotates in the direction where low frictional losses aredesired in step (110), the method generates rotation in the joint withrelatively low friction. The joint turns in the direction where therelative motion of the pair of inclined planes reduces the total columnheight in step (120). The inclined planes are fully collapsed in step(130) so that the normal force between the clutch plate and the pressureplate is determined by the nominal preload generated by the firstpreload spring and the second preload spring. The flat back sides to theinclined planes collide in step (140), causing the flexible element todrive the clutch plate in rotation about the pin in step (150). Frictionexists in the contact between the clutch plate and the pressure plate,but the torque generated by the friction is relatively low due to thelow normal force (160). The pitch of the inclined planes and thefriction properties of the clutch and inclined planes are preferablydesigned to minimize the tendency of the inclined planes to stick whenthe direction of rotation changes from high desired friction to lowdesired friction.

When the joint rotates in the direction where high frictional losses aredesired in step (210), the method generates rotation in the joint withrelatively high friction. The joint turns in the direction where therelative motion of the pair of inclined planes increases the totalcolumn height in step (220). When the compression force on the hinge isless than the critical, the inclined planes rotate and spread apartrather than the clutch plate rotating relative to the pressure platewithout spreading the pair of inclined planes. After a small rotation,the inclined planes increase the total column height in step (230),until all joint end play is removed in step (240), and the column bindsagainst the ends of the joint pin, loading the pin directly in tension.Further rotation causes the inclined planes and the clutch to act as aself-energizing clutch where great normal forces are generated in step(250). The high normal forces caused by the self-energizing clutchresults in a relatively high frictional torque on the joint in step(260). If the torque on the hinge is further increased, the frictionhinge reaches a critical torque in step (270), where the torque to slipthe clutch equals the torque to slip the inclined planes. At torquesabove the critical torque, the clutch slips in step (280). Theself-energizing clutch provides a torque bias, or a large torque orforce difference (preferably a factor of 10 to 12) between rotations inone direction compared to rotations in the opposite direction. Thevalues for the torque bias may be tuned based on the specificapplication and available materials.

Preferred parameters for a friction hinge of the present invention maybe determined based on the force and moment balances on the inclinedplanes and on the clutch as described below. In a preferred embodimentof the present invention, the friction hinge is designed using thefollowing force and moment equations, where:

-   -   T_(A) is the torque applied by the tensioner to the inclined        planes    -   R_(R) is the active radius of the inclined planes    -   R_(C) is the active radius of the clutch    -   μ_(R) is the coefficient of friction of the inclined planes    -   μ_(C) is the coefficient of friction of the clutch    -   N_(R) is the inclined plane normal force    -   N_(C) is the clutch normal force    -   F_(P) is the force from the pin after binding    -   F_(S1) is the arm preload force    -   F_(S2) is the clutch preload force    -   θ is the incline angle of the inclined planes with respect to a        plane parallel to the clutch plate surface

Referring to FIG. 12, the above-listed variables are shown in a freebody diagram during compression for a tensioner arm (91) and a clutch(96) of the present invention.

The arm preload force (F_(S1)) and the clutch preload force (F_(S2)) aredefined as follows, depending on whether the second preload springpresses against the pin, as in FIG. 4, or against the tensioner arm, asin FIG. 2A, FIG. 2B, and FIG. 9. The arm preload force is the netpreload force biasing the tensioner arm against the clutch plate. Theclutch preload force is the net preload force biasing the clutch plateagainst the pressure plate. In the configuration where the secondpreload spring presses against the pin, the arm preload force is thespring force from the first preload spring, and the clutch preload forceis the spring force from the second preload spring.

In the configuration where the second preload spring presses against thetensioner arm, the arm preload force is the difference between thespring force from the first preload spring and the spring force from thesecond preload spring. The spring force from the first spring ispreferably greater than the spring force from the second spring. In thisconfiguration as in the previous configuration, the clutch preload forceis the spring force from the second preload spring.

In compression, the following equations are used:

Pivot Aim Inclined Plane Force Balance:Σ{right arrow over (F)} _(Y)=0=−F _(S1) −F _(P) +N _(R) cos θ−μ_(R) N_(R) sin θ  (1)

Pivot Arm Inclined Plane Moment Balance:Σ{right arrow over (M)} _(O)=0=T _(A) +R _(R) N _(R) sin θ+R _(R)μ_(R) N_(R) cos θ  (2)

Clutch Force Balance:Σ{right arrow over (F)} _(Y)=0=−F _(S2) +N _(C)+μ_(R) N _(R) sin θ−N_(R) cos θ  (3)

Clutch Moment Balance:Σ{right arrow over (M)} _(O)=0=R _(C)μ_(C) N _(C) −R _(R) N _(R) sin θ−R_(R)μ_(R) N _(R) cos θ  (4)

Referring to FIG. 13, the above-listed variables are shown in a freebody diagram during extension for a tensioner arm (91) and a clutch (96)of the present invention.

In extension before the ramps collide, the following equations are used:

Pivot Arm Inclined Plane Force Balance:Σ{right arrow over (F)} _(Y)=0=−F _(S1) −F _(P) +N _(R) cos θ+μ_(R) N_(R) sin θ  (5)

Pivot Arm Inclined Plane Moment Balance:Σ{right arrow over (M)} _(O)=0=T _(A) +R _(R) N _(R) sin θ−R _(R)μ_(R) N_(R) cos θ  (6)

Clutch Force Balance:Σ{right arrow over (F)} _(Y)=0=−F _(S2) +N _(C)−μ_(R) N _(R) sin θ−N_(R) cos θ  (7)

Clutch Moment Balance:Σ{right arrow over (M)} _(O)=0=−R _(R) N _(R) sin θ−R _(C)μ_(C) N _(C)+R _(R)μ_(R) N _(R) cos θ  (8)

During compression prior to pin bind, where F_(P)=0, the clutch must notslip so that the inclined planes climb. In order for the hinge to workas described, the following inequality on the clutch moment must besatisfied:Σ{right arrow over (M)} _(O) =R _(C)μ_(C) N _(C) −R _(R) N _(R) sin θ−R_(R)μ_(R) N _(R) cos θ>0  (4′)

Solving for N_(C):

$\begin{matrix}{N_{C} > \frac{R_{R}{N_{R}\left( {{\sin\;\theta} + {\mu_{R}\cos\;\theta}} \right)}}{R_{C}\mu_{C}}} & \left( 4^{''} \right)\end{matrix}$

Also prior to pin bind, F_(P)=0, so:Σ{right arrow over (F)} _(Y)=0=−F _(S1) +N _(R) cos θ−μ_(R) N _(R) sinθ  (1′)

Solving for N_(R):

$\begin{matrix}{N_{R} = \frac{F_{S\; 1}}{{\cos\;\theta} - {\mu_{R}\sin\;\theta}}} & \left( 1^{''} \right)\end{matrix}$

From the force balance on the clutch:N _(C) =F _(S2)−μ_(R) N _(R) sin θ+N _(R) cos θ  (3′)

Substituting (1″) into (3′):

$\begin{matrix}{N_{C} = {F_{S\; 2} - \frac{\mu_{R}F_{S\; 1}\sin\;\theta}{{\cos\;\theta} - {\mu_{R}\sin\;\theta}} + \frac{F_{S\; 1}\cos\;\theta}{{\cos\;\theta} - {\mu_{R}\sin\;\theta}}}} & (9)\end{matrix}$

Therefore, in order for the inclined planes to climb without clutch slipprior to pin bind, substituting (9) and (1″) into (4″):

$\begin{matrix}{{F_{S\; 2} - \frac{\mu_{R}F_{S\; 1}\sin\;\theta}{{\cos\;\theta} - {\mu_{R}\sin\;\theta}} + \frac{F_{S\; 1}\cos\;\theta}{{\cos\;\theta} - {\mu_{R}\sin\;\theta}}} > \frac{R_{R}{F_{S\; 1}\left( {{\sin\;\theta} + {\mu_{R}\cos\;\theta}} \right)}}{R_{C}{\mu_{C}\left( {{\cos\;\theta} - {\mu_{R}\sin\;\theta}} \right)}}} & (10)\end{matrix}$

Solving for F_(S2)/F_(S1) and simplifying:

$\begin{matrix}{\frac{F_{S\; 2}}{F_{S\; 1}} > \frac{{R_{R}\left( {{\sin\;\theta} + {\mu_{R}\cos\;\theta}} \right)} - {R_{C}{\mu_{C}\left( {{\cos\;\theta} - {\mu_{R}\sin\;\theta}} \right)}}}{R_{C}{\mu_{C}\left( {{\cos\;\theta} - {\mu_{R}\sin\;\theta}} \right)}}} & \left( 10^{\prime} \right)\end{matrix}$

Therefore, the critical ratio of clutch preload to inclined planepreload is a function of the coefficients of friction and the angle ofinclination with respect to a plane parallel to the clutch platesurface.

If R_(R)=R_(C), the relationship can be simplified to:

$\begin{matrix}{\frac{F_{S\; 2}}{F_{S\; 1}} > {\frac{\left( {{\sin\;\theta} + {\mu_{R}\cos\;\theta}} \right)}{\mu_{C}\left( {{\cos\;\theta} - {\mu_{R}\sin\;\theta}} \right)} - 1}} & \left( 10^{''} \right)\end{matrix}$

The critical torque on the friction hinge occurs during compressionafter pin bind at the critical point where the clutch begins to slip.The terms F_(S1), F_(S2), μ_(R), μ_(C), R_(R), R_(C), and θ are knownparameters for a given friction hinge, and the system of four equations(1), (2), (3), (4) is solved for T_(A), N_(R), N_(C), and F_(P). Thusthe performance of the friction hinge may be fine-tuned for a particularapplication by selecting the appropriate forces of the preload springs,the frictional coefficients, the active radii of the hinge, and theincline angle of the planes.

A friction hinge of the present invention may operate in three differentmodes during tensioner compression depending on the torque being placedon it. Referring to FIG. 14, the torque versus arm displacement diagramillustrates these different modes. The torque-arm displacement curve forthe inclined planes (302), (308), (314) and the clutch (304), (310),(316) are shown schematically. The spring preload on the friction hingeis designed such that prior to pin bind, the torque (302) required toturn the inclined planes is less than the torque (304) to turn theclutch. At pin bind (306) the slopes of both the inclined plane curveand the clutch curve increase. The inclined plane torque line (308) hasa slope greater than the slope of the clutch torque line (310) such thatat a critical torque (312), the two lines cross. This is the point atwhich the clutch begins to slip. At torques above the critical point(312), the inclined plane torque line (314) is higher than the clutchtorque line (316), and the clutch slips.

In tensioner compression, the friction hinge follows the path of line(302) to line (308) to line (318). In the first mode, at low compressiontorques prior to pin bind (302), the tensioner arm turns relativelyeasily. This torque is referred to below as “unbound torque”. In thesecond compression mode, at higher compression torques (308) after thepin binds, more torque is required to turn the tensioner arm a givenamount. This torque is referred to below as “bound torque”. In the thirdcompression mode, at very high torques (316) above the critical torque,the clutch slips rather than the inclined planes. The critical torque(312) represents the maximum friction torque that the joint cangenerate, and further arm displacement produces the horizontal torquecurve (318) past the critical point (312). Below the critical point(312) all of the lines are sloped, because the normal force isincreasing due to the elastic deformation of the preload springs or thepin as the stack height increases. After the clutch slips, the columnheight does not increase, so the normal force does not increase.Therefore, the torque generated as the clutch slips is constant.

Following a maximum torque compression event, tensioner extension may gothrough three different modes. The torque follows a curve similar tocurve (308) until the pin is no longer bound. After the pin is unloaded,the torque follows a less steep curve similar to curve (302) until theback sides of the ramps collide. The actual torque values are differentin extension than in compression, because the direction of the rampfriction force changes as the direction of rotation changes. After thecollision the clutch slips, as the clutch plate and tensioner arm at thejoint rotate together. The final torque curve is horizontal, becausethere is no decrease in stack height as the joint rotates after theplanes collide. The normal force remains constant as the clutch slips.

Table 1 shows torque values calculated using equations 1-12 for threedifferent sets of parameters. These values are examples only and serveto illustrate the relationship between the parameters of a frictionhinge and its behavior. Any set of parameters that satisfies thepreviously-described conditions may be used within the spirit of thepresent invention. As a sign convention for torque, negative torquesresist tensioner compression, and positive torques resist tensionerextension. In each of these cases, the inclined planes slip duringcompression at low torques. Calculated outputs include the unboundtorque in compression prior to pin bind, the bound torque in compressionto turn the clutch after pin bind (critical torque), the normal force tobind the pin (F_(P)), the ratio of bound to unbound torque duringcompression, the torque during extension prior to wall collision, thetorque during extension after wall collision, the maximum extensiontorque to unstick the column when the pin is bound, the torque biasbetween compression and extension for large displacements (boundcompression torque/extension torque after collision), the torque biasbetween compression and extension for small displacements (unboundcompression torque/extension torque prior to collision), and the stickbias (full-bound compression torque/unstick torque). The stick biasindicates the likelihood of the inclined planes to stick upon switchingfrom bound tensioner compression to tensioner extension with a largerstick bias indicating a lower stick likelihood. T_(C) values are torquevalues for compression, and T_(E) values are torque values forextension.

TABLE 1 Case 1 Case 2 Case 3 Parameters θ (degrees) 1.0 3.0 5.0 μ_(R)0.1 0.2 0.1 μ_(C) 0.1 0.2 0.2 F_(S1) (N) 1.0 1.0 1.0 F_(S2) (N) 10 15 15R_(R) (mm) 10 10 10 R_(C) (mm) 10 10 10 Calculated Values T_(C) (N · mm)before pin bind −1.18 −2.55 −2.93 T_(C) (N · mm) after pin bind −66.6−138.9 −94.8 F_(P) (N) 55.6 53.5 31.4 T_(C, bound)/T_(C, unbound) 56.654.5 32.4 T_(E) (N · mm) before collide 0.824 1.46 1.11 T_(E) (N · mm)after collide 11 32 32 T_(E) (N · mm) after pin bind 46.7 79.6 35.8Torque bias (large displacements) 6.06 4.34 2.96 Torque bias (smalldisplacements) 1.43 1.75 2.65 Stick bias 1.43 1.75 2.65

Referring to FIG. 15, a two-way clutch of the present invention mayinclude only one preload spring. The friction joint (402) is used at thepivot end (403) of the tensioner. The pivoting proximal end (403) ispivotally mounted to the ground (413) (i.e., the engine housing or abracket) by a pin (414). A mating set of inclined planes (404A), (404B),(405A), (405B) are arranged on two semi-circular contact areas, althoughany number of inclined planes may be used with the tensioner.

Both sets of inclined plane contact areas are centered about a pivothole (408). One set extends from a tensioner arm (418), while the secondset extends from the clutch plate (409). The slope of the inclinedplanes (404A), (404B), (405A), (405B) runs tangential to a constantradius from the center point of the pivot hole (408). The dashed area(488) in FIG. 2A is preferably made of a continuous polymer or ametallic insert molded into the polymer portion of the tensioner arm.The clutch plate (409) has a clutch contact area (410) on the side ofthe cylinder opposite the inclined plane contact areas. The clutchcontact area (410) optionally has friction material (411) bonded to itssurface. The clutch plate (409) mates with a pressure plate (412).

A spring (420) is included to provide a plane gap (415) between theinclined planes (404), (405) and to maintain a nominal preload on theclutch plate (409). The joint is preferably sized so that a smallrelative rotation of the pair of inclined planes (404), (405) takes upthe plane gap (415) and forces the stack of components to contact anddirectly load the pin in tension. The coefficient of friction betweenthe clutch plate (409) and the pressure plate (412) and the coefficientof friction between the mating inclined plane surfaces together with theangle of inclination of the planes, the spring preload, and the activeradii of the clutch and the inclined planes help to determine theperformance of the friction hinge.

The friction hinges described and shown in FIG. 1 through FIG. 15 act astwo-way type clutches, but a friction hinge of the present invention mayalso be a one-way clutch. FIGS. 16 and 17 show a one-way embodiment of apivot joint (502) for a tensioner arm (504). The tensioner arm (504) isbiased toward the tensioned device (506) by a tensioner arm biasingdevice (508). The tensioner arm (504) pivots around a pin (510). The pin(510) is preferably grounded in a pressure plate (512). Stacked on thepin (510) above the pressure plate (512) are a clutch plate (514), thetensioner arm (504), and preferably a washer (516) and a spring (518) tobias the tensioner arm (504) toward the clutch plate (514) and providean end gap (515). The spring (518) is preferably a Belleville spring.The tensioner arm (504) and the clutch plate (514) have complementaryinclined planes (520A), (520B), (522A), (522B) forming complementarycontacting surfaces on the sides facing each other. Although only twosets of inclined planes are shown in FIG. 17, any number of inclinedplanes may be used in the present invention.

A one-way clutch friction hinge of the present invention differs instructure from a two-way friction hinge of the present invention in thatthe one-way hinge does not have a second spring for preloading theclutch plate. A one-way clutch may be used with a single arm tensioneras shown in FIG. 1 or a dual arm tensioner as shown in FIG. 3. Thespring (518) may be a coil spring as shown in FIG. 4 or a torsionspring.

Although the one-way clutch hinge is similar in structure with thetwo-way clutch hinge, it functions differently. The one-way clutch hingeis designed such that the inclined planes slip more easily than theclutch in compression. Since a second preload spring is absent in thisembodiment, slippage of the clutch in compression may be prevented by anincreased coefficient of friction for the clutch or a decreasedcoefficient of friction for the inclined planes. As the torque on thehinge increases, the inclined planes continue to slip and the pinbecomes bound. In this embodiment, as the compression torque continuesto increase, the inclined planes continue to slip until the part fails.The torque to slip the clutch does not become less than the torque toslip the inclined planes in compression, so the clutch does not slip incompression.

For the double spring device, the clutch may slip for an infinitedisplacement in the high torque direction at the maximum torque value.In contrast, the single spring device has a limited possibledisplacement in the high torque direction, because the clutch neverslips in that direction. However, both devices may slip for infinitedisplacements in the low torque direction because the back faces of theramps collide and allow the clutch to be driven. The rotation in thehigh torque direction required for the single spring device to bind thepin and generate enough torque to prevent rotation serves as thebacklash for the one way clutch device.

Accordingly, it is to be understood that the embodiments of theinvention herein described are merely illustrative of the application ofthe principles of the invention. Reference herein to details of theillustrated embodiments is not intended to limit the scope of theclaims, which themselves recite those features regarded as essential tothe invention.

1. A pivot joint comprising: a pressure plate having a pressure platesurface; a clutch plate having a clutch plate hole extending through theclutch plate, a clutch plate surface in contact with the pressure platesurface, and at least one inclined clutch contact area opposite theclutch plate surface; a pivot arm having a pivot hole and comprising atleast one inclined arm contact area in contact with and complementary inshape to the inclined clutch contact area; a pivot pin comprising apivot pin head and a pivot pin shaft extending through the pivot holeand the clutch plate hole and into the pressure plate; a first springmounted on the pivot pin between the pivot pin head and the pivot arm,wherein the first spring urges the pivot arm away from the pivot pinhead, provides a first spring preload on the pivot arm and the clutchplate, and provides an end gap between the pivot pin head and the pivotarm to be taken up by a relative rotation around the pivot pin of the atleast one inclined clutch contact area and of the at least one inclinedarm contact area; and a second spring mounted on the pivot pin andacting on a surface of the clutch plate to bias the clutch plate towardthe pressure plate and providing a second spring preload on the clutchplate; such that when the end gap has not been taken up and the pivotarm rotates in a compression direction, the inclined arm contact arearotates around the pivot pin relative to the inclined clutch contactarea, thereby increasing a distance between the inclined arm contactarea and the pressure plate and increasing a stack height of the clutchplate and the pivot arm; and such that when the end gap has been takenup, the pivot arm binds against the pivot pin, loading the pivot pindirectly in tension.
 2. The pivot joint of claim 1, wherein acoefficient of friction between the clutch plate and the pressure plate,a coefficient of friction between the inclined clutch contact area andthe inclined arm contact area, an angle of inclination of the inclinedarm contact area with respect to a plane parallel to the clutch platesurface, the first spring preload, the second spring preload, an activeradius of the inclined arm contact area, and an active radius of theclutch plate are selected such that in pivot arm compression a torqueload to slip the inclined plane contact area becomes greater than atorque load to slip the clutch plate above a critical torque load thatoccurs at a pivot arm displacement greater than a displacement requiredto take up the end gap.
 3. The pivot joint of claim 1, wherein acoefficient of friction between the clutch plate and the pressure plate,a coefficient of friction between the inclined clutch contact area andthe inclined arm contact area, an angle of inclination of the inclinedarm contact area with respect to a plane parallel to the clutch platesurface, the first spring preload, the second spring preload, an activeradius of the inclined arm contact area, and an active radius of theclutch plate are selected such that in pivot arm extension a torque loadto slip the inclined plane contact area is less than a torque load toslip the clutch plate.
 4. The pivot joint of claim 1, wherein africtional material is located on the clutch plate surface or thepressure plate surface.
 5. The pivot joint of claim 1, wherein the firstspring is a Belleville spring.
 6. The pivot joint of claim 1, whereinthe inclined clutch contact area is centered about the pivot hole andhas a slope running tangential to a constant radius from a center pointof the pivot hole.
 7. The pivot joint of claim 1, wherein the pressureplate is mounted to a stationary surface to prevent rotation of thepressure plate relative to the surface.
 8. The pivot joint of claim 1,wherein the second spring acts on a surface of the pivot arm to bias theclutch plate toward the pressure plate.
 9. The pivot joint of claim 1,wherein the second spring acts on a surface of the pivot pin to bias theclutch plate toward the pressure plate.
 10. The pivot joint of claim 1,wherein a coefficient of friction between the clutch plate and thepressure plate (μ_(C)), a coefficient of friction between the inclinedclutch contact area and the inclined arm contact area (μ_(R)), an angleof inclination of the inclined arm contact area with respect to a planeparallel to the clutch plate surface (θ), an arm preload force (F_(S1)),a clutch preload force (F_(S2)), an active radius of the inclined armcontact area (R_(R)), and an active radius of the clutch plate (R_(C))are selected such that:$\frac{F_{S\; 2}}{F_{S\; 1}} > {\frac{{R_{R}\left( {{\sin\;\theta} + {\mu_{R}\cos\;\theta}} \right)} - {R_{C}{\mu_{C}\left( {{\cos\;\theta} - {\mu_{R}\sin\;\theta}} \right)}}}{R_{C}{\mu_{C}\left( {{\cos\;\theta} - {\mu_{R}\sin\;\theta}} \right)}}.}$11. A tensioner comprising: a pressure plate having a pressure platesurface; a clutch plate having a clutch plate hole extending through theclutch plate, a clutch plate surface in contact with the pressure platesurface, and at least one inclined clutch contact area opposite theclutch plate surface; a tensioner arm for tensioning a chain or belt,the tensioner arm having a pivot hole at a proximal end portion andcomprising at least one inclined arm contact area in contact with andcomplementary in shape to the inclined clutch contact area; a pivot pincomprising a pivot pin head and a pivot pin shaft extending through thepivot hole and the clutch plate hole and into the pressure plate; afirst spring mounted on the pivot pin between the pivot pin head and thetensioner arm, wherein the first spring urges the tensioner arm awayfrom the pivot pin head, provides a first spring preload on thetensioner arm and the clutch plate, and provides an end gap between thepivot pin head and the tensioner arm to be taken up upon tensioning by arelative rotation around the pivot pin of the at least one inclinedclutch contact area and of the at least one inclined arm contact area;and a second spring mounted on the pivot pin and acting on a surface ofthe clutch plate to bias the clutch plate toward the pressure plate forproviding a second spring preload on the clutch plate; such that whenthe end gap has not been taken up and the tensioner arm rotates in acompression direction, the inclined arm contact area rotates around thepivot pin relative to the inclined clutch contact area, therebyincreasing a distance between the inclined arm contact area and thepressure plate and increasing a stack height of the clutch plate and thetensioner arm; and such that when the end gap has been taken up, thetensioner arm binds against the pivot pin, loading the pivot pindirectly in tension.
 12. The tensioner of claim 11, wherein acoefficient of friction between the clutch plate and the pressure plate,a coefficient of friction between the inclined clutch contact area andthe inclined arm contact area, an angle of inclination of the inclinedarm contact area with respect to a plane parallel to the clutch platesurface, the first spring preload, the second spring preload, an activeradius of the inclined arm contact area, and an active radius of theclutch plate are selected such that in tensioner compression a torqueload to slip the inclined plane contact area becomes greater than atorque load to slip the clutch plate above a critical torque load thatoccurs at a tensioner displacement greater than a displacement requiredto take up the end gap.
 13. The tensioner of claim 11, wherein acoefficient of friction between the clutch plate and the pressure plate,a coefficient of friction between the inclined clutch contact area andthe inclined arm contact area, an angle of inclination of the inclinedarm contact area with respect to a plane parallel to the clutch platesurface, the first spring preload, the second spring preload, an activeradius of the inclined arm contact area, and an active radius of theclutch plate are selected such that in tensioner extension a torque loadto slip the inclined plane contact area is less than a torque load toslip the clutch plate.
 14. The tensioner of claim 11, wherein africtional material is located on the clutch plate surface or thepressure plate surface.
 15. The tensioner of claim 11, wherein the firstspring is a Belleville spring.
 16. The tensioner of claim 11, whereinthe inclined clutch contact area is centered about the pivot hole andhas a slope running tangential to a constant radius from a center pointof the pivot hole.
 17. The tensioner of claim 11, wherein the pressureplate is mounted to a stationary surface to prevent rotation of thepressure plate relative to the surface.
 18. The tensioner of claim 11further comprising: at least one blade spring mounted in the tensionerarm; and a distal end portion opposite the proximal end portion; whereinthe distal end portion is slidingly received on a sliding surface. 19.The tensioner of claim 11, wherein a distal end portion opposite theproximal end portion of the tensioner arm is pivotally attached to asecond pivot pin, the tensioner further comprising: at least one bladespring mounted in the tensioner arm; and a second tensioner arm having aproximal end portion pivotally attached to the pivot pin, a distal endportion pivotally attached to the second pivot pin, and a wear surfaceslidingly received on a sliding surface.
 20. The tensioner of claim 11,wherein a portion of the second tensioner arm serves as the pressureplate.
 21. The tensioner of claim 11, wherein a coefficient of frictionbetween the clutch plate and the pressure plate (μ_(C)), a coefficientof friction between the inclined clutch contact area and the inclinedarm contact area (μ_(R)), an angle of inclination of the inclined armcontact area with respect to a plane parallel to the clutch platesurface (θ), an arm preload force (F_(S1)), a clutch preload force(F_(S2)), an active radius of the inclined arm contact area (R_(R)), andan active radius of the clutch plate (R_(C)) are selected such that:$\frac{F_{S\; 2}}{F_{S\; 1}} > {\frac{{R_{R}\left( {{\sin\;\theta} + {\mu_{R}\cos\;\theta}} \right)} - {R_{C}{\mu_{C}\left( {{\cos\;\theta} - {\mu_{R}\sin\;\theta}} \right)}}}{R_{C}{\mu_{C}\left( {{\cos\;\theta} - {\mu_{R}\sin\;\theta}} \right)}}.}$